The present invention relates to a hydraulic drive system for a construction machine including a swing control system, such as a hydraulic excavator. More particularly, the present invention relates to a hydraulic drive system wherein, when a hydraulic fluid from a hydraulic pump is supplied to a plurality of actuators, including a swing motor, through respective associated directional control valves, a delivery rate of the hydraulic pump is controlled by a load sensing system and differential pressures across the directional control valves are controlled by respective associated pressure compensating valves.
JP, A, 60-11706 discloses a hydraulic drive system for controlling a delivery rate of a hydraulic pump by a load sensing system (hereinafter referred to also as an LS system). Also, JP, A, 10-37907 discloses a hydraulic drive system for a construction machine including a swing control system, the hydraulic drive system including an LS system and being intended to realize independence and operability of the swing control system. A 3-pump system mounted on an actual machine is also disclosed as an open-center hydraulic drive system for a construction machine including a swing control system, the hydraulic drive system being intended to realize independence of the swing control system. Further, JP, A, 10-89304 discloses a hydraulic drive system wherein a delivery rate of a hydraulic pump is controlled by an LS system and a pressure compensating valve is given a load dependent characteristic.
In the hydraulic drive system disclosed in JP, A, 60-11706, a plurality of pressure compensating valves each include means for setting, as a target compensation differential pressure, a differential pressure between a delivery pressure of the hydraulic pump and a maximum load pressure among a plurality of actuators. In the combined operation where a plurality of actuators are simultaneously driven, there may occur a saturation state that the delivery rate of the hydraulic pump is not enough to supply flow rates demanded by a plurality of directional control valves. In such a saturation state, the differential pressure between the delivery pressure of the hydraulic pump and the maximum load pressure is lowered, and correspondingly the target compensation differential pressure of each pressure compensating valve is reduced. As a result, the delivery pressure of the hydraulic pump can be distributed again in accordance with a ratio between the respective flow rates demanded by the actuators.
In the hydraulic drive system disclosed in JP, A, 10-37907 and the 3-pump system mounted on an actual machine, an independent open-center circuit using an independent hydraulic pump is constructed for a swing section, which includes a swing motor, separately from a circuit for the other actuators, whereby independence and operability of the swing control system is ensured.
In the hydraulic drive system disclosed in JP, A, 10-89304, a plurality of pressure compensating valves each have hydraulic pressure chambers constructed as follows. A pressure bearing area of a hydraulic pressure chamber, to which an input side pressure of a directional control valve is introduced and which produces a force acting in the valve-closing direction, is set to be greater than a pressure bearing area of a hydraulic pressure chamber, to which an output side pressure of the directional control valve is introduced and which produces a force acting in the valve-opening direction. The pressure compensating valve is given such a load dependent characteristic that, as a load pressure of each associated actuator rises, the target compensation differential pressure of the pressure compensating valve is reduced (i.e., the pressure compensating valve is throttled) to decrease a supply flow rate to the actuator. As a result, the actuators on both the lower and higher load sides can be operated with good operability in a stable manner without hunting. Further, a ratio of the pressure bearing area of the hydraulic pressure chamber, to which the output side pressure of the directional control valve is introduced, to the pressure bearing area of a hydraulic pressure chamber, to which the input side pressure of the directional control valve is introduced, is specified to fall in the range of 0.97-0.94.
The conventional hydraulic drive systems described above however have the following problems with the swing control system.
JP, A, 60-11706: following problems {circle around (1)} and {circle around (2)}
JP, A, 10-89304: following problem {circle around (2)}
JP, A, 10-37907: following problem {circle around (3)}
Open-center 3-pump system mounted on actual machine: following problem {circle around (3)}
{circle around (1)} jerky feel in operation at start-up of swing
{circle around (2)} occurrence of energy loss, vibration, heat, etc. at start-up of swing
{circle around (3)} increase in cost and space and complicated circuit configuration due to provision of a separate circuit
(1) JP, A, 60-11706
When the hydraulic drive system including the LS system, disclosed in JP, A, 60-11706, is applied to the swing control system, it is difficult to keep balance between load sensing control (hereinafter referred to also as an LS control) of the hydraulic pump and a flow rate compensating function of the pressure compensating valve due to an inertial load of the swing control system. This is because a difficulty occurs in keeping balance between response of the pressure compensating valve and response of the LS control for the hydraulic pump due to the following reasons when a swing driving pressure is controlled in a stage of shift from swing acceleration to steady rotation.
(1) In a swing start-up and acceleration mode, the pump LS control is performed so as to raise a delivery pressure of the hydraulic pump depending on the swing start-up pressure for holding a constant flow rate.
(2) To hold constant a differential pressure across a throttling element of the directional control valve, the pressure compensating valve is operated in a direction to increase a flow rate passing itself that tends to reduce upon a rise of the load pressure.
(3) When the swing reaches a steady speed, the swing driving pressure is lowered and therefore the pump LS control is not required to control the delivery pressure of the hydraulic pump so high as in the swing start-up and acceleration mode. Hence the pump LS control is performed in a direction to lower the delivery pressure of the hydraulic pump.
(4) Upon a lowering of the swing driving pressure, the pressure compensating valve is operated in a direction to reduce the flow rate passing itself that tends to increase.
Because of quick shift from (1) to (4), the swing operation becomes jerky (above problem {circle around (1)}).
Further, in the above steps (1) and (2) of the swing start-up and acceleration mode, the hydraulic fluid is supplied to the swing motor at a flow rate larger than a necessary level. As a result, the load pressure of the swing motor rises to a pressure set by an overload relief valve that serves as a swing safety valve, and a large amount of hydraulic fluid corresponding to an extra flow rate is drained to a reservoir through the swing safety valve. The extra flow rate results in energy loss, thereby deteriorating energy efficiency, and also gives rise to vibration, heat and noise (above problem {circle around (2)}).
(2) JP, A, 10-89304
In the hydraulic drive system disclosed in JP, A, 10-89304, since the pressure compensating valve is given a load dependent characteristic, the target compensation differential pressure of the pressure compensating valve is reduced in response to a rise of the load pressure of the swing motor when swing is solely started up, and when the swing motor shifts to a steady sate, the target compensation differential pressure of the pressure compensating valve is also returned to the original value in response to a lowering of the load pressure of the swing motor. As a result, swing can be started up without causing a jerky feel in operation. To provide the load dependent characteristic, however, the pressure bearing area ratio is specified to fall in the range of 0.97-0.94. By setting the pressure bearing area ratio in such a way, the proper load dependent characteristic is not always provided for all of different machine specifications (such as inertial load, swing device capacity, supply flow rate, and swing angular speed). In the swing start-up and acceleration mode, therefore, the swing motor is supplied with the hydraulic fluid at a considerable extra flow rate and a substantial amount of hydraulic fluid corresponding to the extra flow rate is likewise drained to a reservoir through a swing safety valve. As with the above case, the extra flow rate results in energy loss, thereby deteriorating energy efficiency, and also gives rise to vibration, heat and noise (above problem {circle around (2)}).
(3) Hydraulic Drive System Disclosed in JP, A, 10-37907 and Open-center 3-Pump System Mounted on Actual Machine
In the hydraulic drive system disclosed in JP, A, 10-37907, the swing control system is constructed by a separate open-center circuit to ensure satisfactory swing operability in the LS system. Also, in the open-center 3-pump system mounted on an actual machine, the swing control system is constructed as a separate open-center circuit to ensure satisfactory swing operability.
More specifically, in the open-center system, when the driving pressure rises at the swing start-up, a flow rate of the hydraulic fluid returning to a reservoir through a center bypass fluid line is increased, which reduces a flow rate of the hydraulic fluid passing a throttle of the directional control valve for the swing section. A flow rate of the hydraulic fluid supplied to the swing motor is therefore restricted in the swing start-up and acceleration mode. When the swing speed reaches a steady speed, no restriction is imposed on the supply flow rate to the swing motor because of the driving pressure being not so high as at the swing start-up, and the hydraulic fluid is supplied to the swing motor at a flow rate corresponding to an opening of the throttle of the directional control valve for the swing section. The swing can be thereby smoothly started up without causing a jerky feel in operation for starting up the swing solely unlike the LS control. Also, the hydraulic fluid is suppressed from being supplied to the swing motor at an extra flow rate larger than a necessary level. In the combined operation of the swing motor and any other actuator, therefore, a part of the delivery rate of the hydraulic pump, which is saved from being supplied to the swing motor, can be supplied to the other actuator, thus resulting in more efficient and stable operation.
However, in the hydraulic drive system disclosed in JP, A, 10-37907 and the 3-pump system mounted on an actual machine, the swing control system must be constructed as a separate circuit in parallel to the system for the other actuators. Correspondingly, a cost is pushed up and a space required for installation is increased. In addition, a hydraulic pump for the swing control system must be separately provided. In the system disclosed in JP, A, 10-37907, particularly, a signal line is required to keep power balance between the swing control system and the LS system which are arranged in parallel, and hence a circuit configuration is complicated (problem {circle around (3)}).
An object of the present invention is to provide a hydraulic drive system including a swing control system, which enables swing operation to be accelerated for shift to a steady state without causing a jerky feel at the start-up of swing, which can realize a stable swing system with good energy efficiency, and which is free from problems resulted from providing a separate circuit, such as an increase in cost and space and complication of a circuit configuration.
(1) To achieve the above object, the present invention provides a hydraulic drive system comprising a hydraulic pump, a plurality of actuators, including a swing motor, which are driven by a hydraulic fluid delivered from the hydraulic pump, a plurality of directional control valves for controlling respective flow rates of the hydraulic fluid supplied from the hydraulic pump to the plurality of actuators, a plurality of pressure compensating valves for controlling respective differential pressures across the plurality of directional control valves, and pump control means for load sensing control to control a pump delivery rate such that a delivery pressure of the hydraulic pump is held a predetermined value higher than a maximum load pressure among the plurality of actuators, wherein the hydraulic drive system further comprises target compensation differential-pressure setting means provided respectively in the plurality of pressure compensating valves and setting, as a target compensation differential pressure, a differential pressure between the delivery pressure of the hydraulic pump and the maximum load pressure among the plurality of actuators, and target compensation differential-pressure modifying means provided in the pressure compensating valve of the plurality of pressure compensating valves, which is associated with a swing section including the swing motor, for giving the pressure compensating valve for the swing section such a load dependent characteristic that when the load pressure of the swing motor rises, the target compensation differential pressure of the pressure compensating valve for the swing section, which is set by the target compensation differential-pressure setting means, is reduced to provide a flow rate characteristic simulating constant-horsepower control of the swing motor.
By thus providing the target compensation differential-pressure modifying means in the pressure compensating valve for the swing section and giving the pressure compensating valve for the swing section the load dependent characteristic, the pressure compensating valve for the swing section performs fine adjustment of a flow rate in accordance with change in the load pressure of the swing motor at the swing start-up so that the swing motor is smoothly accelerated for shift to a steady state.
Also, by giving the pressure compensating valve for the swing section the load dependent characteristic that provides the flow rate characteristic simulating the constant-horsepower control, it is possible to carry out control such that energy per unit time supplied to the swing motor in a start-up and acceleration mode approximates an energy value in the steady state to be eventually reached. At transition from the start-up and acceleration mode to the steady state, therefore, energy required for accelerating a swing structure is ensured so as to maintain accelerating performance (acceleration feel) and useless energy is not supplied to the swing motor. Accordingly, an extra flow rate of the hydraulic fluid drained to the reservoir through an overload relief valve is reduced, whereby a stable swing system with good energy efficiency can be constructed.
Additionally, since the above-described function can be achieved without providing a separate circuit, problems of an increase in cost and space and complicated circuit configuration are avoided.
(2) In the above (1), preferably, the flow rate characteristic simulating the constant-horsepower control is such a characteristic that a flow rate resulted at the load pressure immediately after the start-up of the swing motor is substantially equal to a flow rate providing a horsepower equal to the horsepower outputted in a steady state of the swing motor.
With that feature, energy per unit time supplied to the swing motor is controlled so as to approximate an energy value in the steady state from immediately after the start-up, and satisfactory accelerating performance is achieved while a stable swing system with good energy efficiency is constructed.
(3) Also, in the above (1), preferably, the flow rate characteristic simulating the constant-horsepower control is such a characteristic that a flow rate resulted at the load pressure immediately after the start-up of the swing motor is substantially equal to a flow rate in a predetermined range set with a flow rate, as a reference, providing a horsepower equal to the horsepower outputted in a steady state of the swing motor.
With that feature, energy per unit time supplied to the swing motor is controlled so as to approximate an energy value in the steady state from immediately after the start-up, and satisfactory accelerating performance is achieved while a stable swing system with good energy efficiency is constructed.
(4) In the above (3), for example, the flow rate characteristic simulating the constant-horsepower control may be such a characteristic that a flow rate resulted at a load pressure, which is substantially middle between the load pressure in the steady state and the load pressure immediately after the start-up, is not smaller than a flow rate providing a horsepower equal to the horsepower outputted in the steady state of the swing motor.
With that feature, an extra flow rate of the hydraulic fluid drained to a reservoir is reduced with a decrease in the flow rate resulted at the load pressure immediately after the start-up of the swing motor. Therefore, effects of improving energy efficiency and enhancing stability are further increased.
Also, since the flow rate resulted at the load pressure, which is substantially middle between the load pressure in the steady state and the load pressure immediately after the start-up, is not smaller than the flow rate providing a horsepower equal to the horsepower outputted in the steady state of the swing motor, accelerating performance can be ensured.
(5) Further, in the above (1) to (3), preferably, the pressure compensating valve for the swing section has signal pressure bearing chambers on which an input sisde pressure and an output side pressure of the directional control valve for the same swing section act respectively as signal pressures, and the target compensation differential-pressure modifying means is constructed by providing an area difference between the signal pressure bearing chambers of the pressure compensating valve for the swing section and setting a pressure bearing area ratio between the signal pressure bearing chambers to provide the above flow rate characteristic.
With those features, the target compensation differential-pressure modifying means can be constructed in fully hydraulic fashion.
(6) Moreover, in the above (1) to (3), the target compensation differential-pressure modifying means may comprise means for detecting a load pressure of the swing motor, a controller for calculating a target flow rate corresponding to the detected load pressure based on a preset constant-horsepower control characteristic, and outputting a control signal corresponding to the calculated target flow rate, and means operated by the control signal for modifying the target compensation differential pressure of the pressure compensating valve for the swing section so that the target flow rate is obtained.
With those features, the target compensation differential-pressure modifying means can be constructed using a controller.